The present invention relates to a heat exchanger, and particularly to a heat exchanger having a heat transmission element, such as fins, with improved heat transmission characteristics.
FIG. 1a and 1b are a front view and a side view, respectively, of a conventional heat exchanger of the plate fin-tube type, in which reference numeral 1 depicts a plurality of first heat transmission members in the form of parallel fins arranged in a fluid flow direction A, and 2 a plurality of second heat transmission members in the form of parallel pipes whose temperature is different from that of the first heat transmission members 1 and which are thermally connected to the pipes 2 by pressure contact or soldering. A primary fluid flows through the pipes 2 and a secondary fluid flows outside the pipes, i.e., between the fins 1. Heat exchange is performed between the first and second fluids.
FIG. 2a and 2b are a front view and a side view, respectively, of a conventional heat sink of a semiconductor element, which is a type of heat exchanger. In FIGS. 2a and 2b, reference numeral 21 depicts a solid rod which acts as the second heat transmission member and is thermally coupled to the fin 1 by pressure contact or soldering. A semiconductor element (not shown) is pressure contacted to an end face 22 of the solid rod 21. Heat generated in the semiconductor element is transmitted through the solid rod 21 to the fin 1, from which heat is dissipated.
A heat pipe may be used instead of the solid rod 21. The use of a heat pipe is particularly useful when used together with a high performance fin because the heat pipe makes the axial temperature distribution uniform.
In the heat exchanger shown in FIGS. 1a to 2b, the total area of the fins 1 is usually about 20 times the total surface area of the tubes 2 or the solid rod 21, and therefore the heat transmission characteristics of the fins affect the performance of the heat exchanger substantially.
It is assumed for simplicity that the fin 1 is a flat plate having no holes for the pipes 2 or the solid rod 21 since the area to be occupied by these holes is actually very small.
For such a flat fin 1, there have been various methods proposed to improve the heat transmission characteristics, such as making a temperature boundary layer as thin as possible.
Describing the temperature boundary layer, FIG. 3 is a perspective view of a portion of a corrugated fin type heat exchanger, which is widely used in automotive radiators, etc. In FIG. 3, a second heat transmission element, i.e., pipes 2, through which a primary fluid B such as engine coolant flows are thermally connected to a first heat transmission element, i.e., a corrugated fin 1. A second fluid A such as air flows through gaps formed by the corrugations of the fin 1. The corrugated fin 1, which is equivalent to a plurality of parallel flat fins, has a defect which will be described with reference to FIG. 4, which shows an air flow A around a portion of the fin 1 in FIG. 3.
According to the general theory of heat transmission, when coolant air A flows along opposite, surfaces of the fin 1, a temperature boundary layer 3 is produced along air flow A as shown in FIG. 4. The temperature distribution of the air within the boundary layer 3 is shown by a dotted line in FIG. 4, wherein the temperature of the fin wall is indicated by t.sub.w, the temperature of the air flow A outside the boundary layer 3 by t, and the distance from the fin wall by x. The heat transfer coefficient .alpha. from the fin 1 to the air flow A is defined in this case by: ##EQU1## That is, the variation of for a system in which t, t.sub.w and the thermal conductivity k are constants corresponds to (dt/dx)w, i.e., the gradient of the temperature distribution of the air in the vicinity of the surfaces of the fin 1. That is, the heat transfer coefficient is in proportion to the gradient of the temperature distribution of the fluid in contact with the fin surfaces, which in turn is proportional to tan .theta..
Further, since (t.sub.w -t.infin.) is a constant, the thicker the boundary layer 3, the smaller the angle .theta..
Still further, local heat transfer coefficient in the temperature boundary layer 3 produced along the fin 1 are reduced, and thus the average heat transfer coefficient, namely, the average of the local transmittances, is very low.
In order to resolve this problem, various proposals have been made.
An example of one such proposal is shown in FIG. 5, which shows a perspective view of a portion of a heat exchanger of a type widely used in an automotive or aircraft radiator. The heat exchanger shown in FIG. 5 is referred to as being of the "offset fin" type in which the fin 1 is divided into a plurality of fin pieces (referred to as "strips" hereinafter) as shown. With such strips, the temperature boundary layer 3 is also divided as shown in FIG. 6 (corresponding to FIG. 4), and thus the average thickness of the boundary layer is reduced, resulting in a higher average heat transfer coefficient.
This effect, termed a "leading edge" effect, is utilized effectively in various heat exchangers or other heat transmitting equipment. For example, as seen in FIG. 7, the principle is applied to a heat transmission fin of the plate fin-tube type heat exchanger for use in an air-conditioning apparatus. In FIG. 7, a plurality of fins 10 are arranged in parallel and a plurality of heat transmission pipes through which coolant flows are passed through pipe insert holes 12 of the fins 10, extending orthogonally thereto. The fin 10 is partially stepped to form raised strips 11 so that the boundary layer is divided as shown in FIG. 8.
FIG. 9 shows another example of a fin configuration, specifically, of a type disclosed in Japanese Laid-Open Utility Model Application No. 58184/1981, in which strips 11 are formed at an angle to a fin 10 and the secondary fluid A flows along the strips 11. The configuration of the strips provides the leading edge effect.
FIG. 10 shows in plan view another fin configuration, which is disclosed in Sanyo Technical Review, vol. 15, no.1, February 1983, page 76, and FIG. 11 is a cross section taken along a line XXX-XXX in FIG. 10. In these figures, a fin 10 is formed, in an area between adjacent heat transmission pipes 12, with corrugations in each of which pressed-up portions 11 are formed. In this configuration, the fin is divided into a plurality of inverse-V shaped strips so that fluid flow A is deflected thereby.
FIG. 12 shows another example of a conventional fin, specifically, a fin referred to as a louver fin. Coolant A flows between adjacent strips 11 as shown by a dotted line, and thus the leading edge effect is obtained.
FIG. 13 depicts another example, disclosed in Japanese Laid-Open Patent Application No. 105194/1980, in which a main fluid A flows between fins 1a and 1b, each formed with a plurality of slits 13 orthogonal to the fluid flow, while passing through the slits. The leading edge effect is provided by an area between adjacent slits.
Problems inherent commonly to these conventional fin configurations utilizing the leading edge effect will be described with reference to FIG. 6.
Firstly, the pressure loss is increased considerably. That is, a boundary layer 3 is produced for each strip but is broken at the trailing edge of the strip. Then, another boundary layer is produced again at a leading edge of a succeeding strip. When the secondary fluid is air (whose Prandtl number Pr is nearly equal to 1), a temperature boundary layer is analogous to a velocity boundary layer. That is, if the temperature boundary layer is thin, the velocity boundary layer is also thin, meaning that the velocity gradient on the heat transmission surface is increased relatively, resulting in a considerable increase of friction loss. As another source of pressure loss, there is a resistance due to the leading edge configurations of the strip, which has a non-negligible thickness. In addition, generally either or both edges of the strip have flashes formed during the fabrication thereof. Therefore, the increase of resistance due to the strip configuration is usually considerable.
Secondly, the degree of improvement of heat transmittance attributable to the leading edge effect is not so much as desired. Specifically, because there exists a velocity loss area behind each strip, the subsequent strip is influenced by such field of velocity, resulting in a reduction of heat transmittance. The same applies for temperature considerations.
In view of the leading edge effect, the strip should be as narrow as possible. In fact, the heat transmittance is improved if the width of the strip is reduced to some extent. However, if the width of the strip is reduced beyond a certain value, the heat transmittance cannot be improved and may be reduced in some cases. Since the reduced width of the strip means a reduced interval between adjacent strips in the second fluid flow direction, the improvement of heat transmittance may be restricted thereby. The conventional configurations shown in FIG. 9 and 11 are employed to avoid such undesirable effects.
Further, a relative reduction of fin efficiency due to the employment of the divided fins is another reason for the restricted heat transmittance.
It has been empirically concluded that the heat transfer coefficient of the fin utilizing the leading edge effect is increased up to by 50% of that of the flat fin and the pressure loss is about t.sub.w ice that of the latter.
Another problem is the mechanical strength of the fin, which is reduced by increasing the number of strips. This problem has become more severe due to the recent tendency of reducing the thickness of the fins for economic reasons.